Two-stage cooling system for heat machine components



July 2, 1968 C. A. E. BEURTHERET TWO-STAGE COOLING SYSTEM FOR HEATMACHINE COMPONENTS Filed Aug. 23, 1966 4 Sheets-Sheet 1 l/vrewroe Omens45 Bflf/WH July 2, 1968 C. A. E. BEURTHERET TWO-STAGE COOLING SYSTEM FORHEAT MACHINE COMPONENTS Filed Aug. 23, 1966 4 Sheets-Sheet 2 A fig 2 18E 13 3o 1- =-:i: 4 as 32A 6 699% 321 QGGAX 9 1- T- 35A 31 37 15 Q I I'11 3B 40 Z D K: \ac 20 1 PUMP O on.

COOLER y 1968 c. A. E. BEURTHERET 3,390,667

TWO-STAGE COOLING SYSTEM FOR HEAT MACHINE COMPONENTS Filed Aug. 23, 19664 Sheets-Sheet 3 'lll III/ 4 July 2, 1968 c. A. E. BEURi'HERET 3,

TWO-STAGE COOLING SYSTEM FOR HEAT MACHINE COMPONENTS Filed Aug. 25, 19664 Sheets-Sheet 4 Q so United States Patent Olfice 3399,65? Patented July2, 1968 17 Claims. cr. 123-3 ABSTRACT OF THE DISCLOSURE A sealed primarychamber containing a vaporizable liquid such as water is arrangedadjacent the wall of the piston exposed to heat. This liquid issubmitted to stabilized vaporization and condensation, transferring itsheat through a separating Wall to a secondary coolant fluid,specifically lubricant oil, circulated through a secondary, innerchamber of the piston.

This invention relates to cooling systems for the heated components ofheat machines, including both engines and compressors.

The invention has especial utility in connection with the cooling ofpistons of combustion engines, both of the reciprocatory-piston type andthe rotary-piston type. It will, therefore, be described in terms ofsuch use, but this should not be construed as a limitation on the scopeof the invention.

The pistons of combustion engines serve to define variable-volume spacesor compression chambers containing hot gases and/ or combustion mixturesunder high pressures, and the piston must provide a tight seal againstsuch pressures with the help of piston rings or equivalent sealingelements. If the pistons and their seals are to op crate correctly, itis essential that the large amounts of heat generated in the compressionchambers to which piston surfaces are exposed, shall be dissipated at asufiiciently high rate to the exterior.

The problem of providing adequate rates of heat dissipation from thepistons of large-sized engines has been found especially diflicult tosolve in the case of the multilobed rotary pistons, or rotors, of theso-called rotarypiston engines, in which the rotor and stator are formedwith interengaging lobes differing in number by unity, so that the rotorrevolves in a complex path within the stator to define thevariable-volume chambers therein. The path of the rotor within thestator may be epicycloidal (sometimes called epitrochoidal) orhypocycloidal (also called hypotrochoidal) depending on the geometry ofthe machine. Such paths will generically be referred to as trochoidalherein.

The problem of providing an adequate rotor cooling system has been oneof the main causes that has so far held up the practical development ofrotary-piston machines, which otherwise are very attractive and holdgreat promise for the powering of vehicles and other applications.

Since the moving parts of a heat machine must be lubricated, it wouldappear natural to use the lubricant as a coolant to dissipate thegenerated heat. This concept has been Widely utilized in rotary-pistonengines, and many different types of rotary-piston machine constructionshave been disclosed, wherein the lubricant is circulated throughsuitable recesses and compartments formed in the rotor in order to carryaway with it the heat generated adjacent the rotor surfaces exposed tohigh temperatures.

However, the results have been only moderately successful. Lubricantoils are poor heat transfer agents and do not allow a rate of heatdissipation as high as would be desirable. Oils are chemically unstableand break down rapidly at the high temperatures that would beadvantageous in the operation of such machines, requiring frequentlubricant replacement. Further, the high inertial forces generated bythe reciprocatory components of the trochoidal motion of the rotor in arotary-piston machine cause foaming and frothing of the oil andseparation of the liquid from the heated surfaces, thereby furtherreducing heat-transfer efficiency.

It has also been proposed to seal a body of a liquid that is a goodconductor of heat, such as potassium or sodium (which liquefy at thetemperatures involved) in one or more sealed cavities formed in therotor of a rotarypiston engine near the rotor surfaces exposed to hightemperatures, in order to abstract the heat therefrom and transfer it toa secondary coolant such as lubricating oil circulated through therotor. This expedient has made it possible somewhat to improve thecooling of rotor surfaces remote from the axis of rotation, but theimprovement in over-all heat dissipating rate achieved in this way hasbeen slight and has hardly justified the increase in complication andexpense.

The applicant has been engaged for many years in the study ofevaporation cooling processes. Evaporation cooling, in which heat isabstracted from a heated surface by a boiling liquid in contact with thesurface, is very attractive in theory because it should enableattainment of much higher heat dissipation rates than would be possiblewith a single-phase liquid coolant, due to the high latent heats ofvaporization of certain common liquids including water. Early attemptsat evaporation cooling encountered frustrating difiiculties because ofso-called spheroidal-state ebullition and resulting burn-out. That is,it was found that wherever the temperature of the heatdissipatingsurface locally exceeded a rather moderate critical temperature (aboutC. for water at ordinary pressure), spheroidal-state ebullition wouldset in and the temperature would tend to jump almost instantly to anenormously higher value (of the order of 1000 C.), causing destructiveburnout at a point of the metal surface.

In a number of earlier US. patents and other publications, the applicanthas disclosed means for positively avoiding this runaway temperaturecondition and consequent burn-out. The applicants evaporation coolingsystems, known commercially as Vapotron broadly involve the provision ofheat-dissipating formations in the form of protuberances (bosses and/ orribs) separated by channels or grooves, projecting from theheat-dissipating surface into the boiling liquid, and so shaped anddimensioned that stable temperature gradients are established alongtheir sides. These stable temperature gradients which encompass theso-called critical temperature referred to above, are found to be apositive safeguard against runaway temperature and consequent burn-out.The applicants Vapotron technique have thus made it feasible to reachrates of heat dissipation and cooling efficiencies very greatly superiorto those achievable with singlephase liquid cooling. In the past years,Vapotron technique have been successfully applied on a worldwideindustrial scale to the cooling of high-power electron discharge tubes,and have made it possible to construct such tubes having power ratingsvery many times greater than was earlier possible. Vapotron coolers havealso been applied to the cooling of fuel elements in nuclear powerplants, and in other applications.

It has been the applicants main object in the present invention to applythe so-called Vapotron, or non-isotherm evaporation-cooling, techniques,to the moving parts of heat machines, and particularly to the pistons ofrotarypiston and reciprocatory-piston combustion engines.

As indicated earlier herein, lubricant oils are unsuitable for use notonly as single-phase liquid cooling agents but also as two-phaseevaporation-cooling agents because, inter alia, of their chemicalinstability at high temperatures. The evaporating liquid mustnecessarily be a liquid having a boiling point in the range oftemperatures permissible for the part to be cooled, and one that willnot break down in that temperature range. Water satisfies theserequirements. However, the provision of a system for circulating wateror another evaporable liquid, in addition to lubricant, through areciprocatory piston or a rotor, raises very serious design andconstructional problems owing to the sealing requirements, and theresulting system would be unsatisfactory from an engineering standpoint.It has been an object of this invention to eliminate these difficultiesentirely, through the provision of a two-stage cooling system in which abody of evaporable liquid (such as water) is permanently enclosed in asealed chamber of the rotor or other moving part to be cooled, adjacentthe heated surfaces thereof, and in which the sealed evaporable liquidtransfers its heat to a secondary coolant fluid, specifically lubricantoil, circulated through another chamber of the moving part in heattransfer relation with the evaporable liquid.

In applying the above concepts to heat machines in which the movingparts have components of motion subject to periodically varyingaccelerations, as is the case with the pistons and rotors of bothreciprocatingpiston and rotary-piston engines, a further problem isencountered. Due to the inertial forces generated in the sealed body ofprimary evaporable liquid by the varying accelerations, the heated wallsurfaces are intermittently stripped of liquid during certain periods ofthe cyclic motion, and are exposed merely to the vapor so that duringsuch periods heat transfer becomes very poor, and there is introduced adefinite risk of subsequent spheroidalstate ebullition and burn-out. Itis an object of this invention to eliminate this danger entirely, and itis in fact a further object hereof to utilize the cyclic inertial forcescreated by the reversing accelerations of the reciprocatory motion of a,piston (whether in a reciprocatory-piston machine or in a rotary-pistonmachine) actually to improve the efificiency of the heat transferprocess.

The above and further objects of the invention will be made clear fromthe ensuing description of exemplary embodiments with reference to theaccompanying drawings, wherein:

FIG. 1 is a simplified cross sectional view of a rotarypiston engineprovided with cooling means according to the invention, the sectionbeing on the line A-A of FIG. 2;

FIG. 2 is a sectional view on line B-B of FIG. 1;

FIG. 3 is a fragmentary view showing the general conditions of the fluidin the primary chamber during the inward stroke of the rotary piston;

FIG. 4 similarly shows the conditions during the outward stroke of thepiston;

FIG. 5 shows the conditions at an intermediate stage during movementreversal; and

FIG. 6 is a simplified sectional View of a reciprocatingengine pistonprovided with cooling means according to the invention in a differentembodiment.

A rotary-piston engine of generally conventional layout is illustratedin FIG. 1 as comprising a five-lobed stator 1 and an inner four-lobedrotor 5.

The stator 1 includes an outer casing wall 1A of generally pentagonalshape and an inner wall 2 having a contour resembling a five-pointedstar. The inner stator wall 2 internally defines five concaves lobes 2.The meeting of each pair of adjacent lobes defines a cusp at which aninwardly-opening slot 11 is formed, with a seal strip 11' positioned inthe slot. In the central region of each lobe the inner stator wall 2projects outwards to form a suitably shaped compression chamber 19sealed at its outer end which is spaced inward from the outer casingwall 1A so as to define a peripheral space 12 between the inner wall 2and outer wall 1A of the stator. As can be seen from FIG. 2, the outerwall 1A and the inner wall 2 of the stator 1 constitute an integralcasting having parallel spaced transverse end walls 2C and 2D, so thatthe space 12 constitutes a sealed enclosure. The structure thusdescribed is firmly, and preferably dismountably, assembled as by bolts(not shown) between the transverse end flanges 1C and ID of the stator.

The rotor 5 is in the form of a four-lobed recessed shell havingtransverse walls 15 and 16 rotatably slidable in close mating engagementwith the stator sidewalls or end flanges 1C and 1D. Annular sealingstrips 17 inserted in slots formed in the outer surfaces of the rotorsidewalls 15 and 16 near the outer edges thereof engage the innersurfaces of stator sidewalls 1C and 1D. The rotor shell extending acrossthe rotor end walls 15 and 16 is four-lobed as is clearly apparent fromFIG. 1, and each of the four outwardly convex lobes 6, 7, 8 and 9 of therotor is adapted for close mating engagement with each of the fiveconcave lobes of the stator shell 2. It will be understood that in theoperation of the engine only one of the rotor lobes can at anyparticular time be fully seated in a concave stator lobe, and in FIG. 1this condition is shown for the uppermost rotor lobe 6.

The rotor end walls 15 and 16 are provided in their central regions withhubs 18 and 19 respectively, which are secured to said end walls by wayof annular insert members 27 and 28, secured in said end walls by way ofinterfltting shoulder joints and annular seals 29 as shown, forconvenience of assembly.

The hubs 18 and 19 are formed with cylindrical inner bearing surfaceswhich rotatably surround an eccenter member 4. The smooth bearingsdescribed may be replaced with antifriction bearings if desired. Theeccenter 4 is part of the engine output shaft 3, which is rotatablyreceived in cylindrical bearing surfaces formed internally of end hubsformed integrally with the stator sidewall flanges 1C and ID as shown.The smooth stator bearings of the rotor may likewise be replaced withantifriction bearings.

Secured around the rotor hubs 18 and 19 are externally toothed gearannuli 20 and 21, which mesh with internal annuli 22 and 23 secured inshoulders defined within the stator sidewalls 1C and 1D.

Each of the five stator combustion chambers 10 is equipped withconventional fuel injection, ignition, and exhaust means, here shownonly schematically as including an intake valve and an exhaust valve 81and 82 projecting into each compression chamber 10 of the stator, and anignition plug 83 (or a fuel injector, as the case may be) projectingcentrally into the chamber.

The operation of this type of rotary-piston engine is well known in theart and will only be briefly described to the extent that it ispertinent to the invention.

The rotary piston or rotor 5 is constrained to describe a complexmotion, epitrochoidal in character, within the stator, during which eachof the four rotor lobes 6-9 is successively seated in a stator lobe,thereby Sealing the corresponding stator compression chamber 10-. It canbe shown that as the engine shaft 3 rotates, the rotor 5 has a componentof rotation about the center axis C of the shaft, in a direction reversefrom that of the shaft at an angular rate four times slower than that ofthe shaft as indicated by the arrows. At the same time, due to theeccenter 4 the rotor 5 has another component of motion which is acircular translation described bodily with the circular translation ofthe center of said eccenter 4 about the shaft axis C, and this lastcomponent of motion imparts to each of the rotor lobes 6-9 a radialreciprocation toward and away from the center axis C. The result of thiscomposite movement of the rotor is to cause a cyclic variation of theeffective volume of the sealed chambers defined between the rotor andstator. The timing of the intake and exhaust valving and the ignitionmeans is so predetermined as to ignite the combustible fuel-air mixturein each of the five stator chambers 10 in succession,

as said chamber is in the minimum-volume sealed condition shown for theuppermost chamber in FIG. 1. The ignited mixture then expands as thesealed volume is progressively increased, and is exhausted through theexhaust valve 82 of the chamber, after which a fresh amount of fuelmixture is admitted into the chamber by the operation of the intakevalve 81, and is thereafter compressed as the sealed volume diminishes.In the embodiment described the engine thus operates in a fourstrokecycle. It is important to note that in this type of operation, thecoaction of the rotor and stator is such that it is always the same pairof rotor lobes, herein lobes 7 and 9 respectively, which are exposed tothe gases at the intake and exhaust strokes of the cycle, and always theother pair of rotor lobes, as shown the lobes 6 and 8, which are exposedto the gases during compression and expansion. As a consequence, therotor lobes are unequally loaded thermally, the lobes 6 and 8 beingexposed to considerably more heat than the lobes 7 and 9.

It will be understood that the illustrated embodiment is but one exampleof the wide variety of operative rotarypiston engines that can beconstructed, and to all of which the invention is applicable.

The lubricating and cooling system of the improved engine according tothe invention will now be described.

The inner space of the recessed rotor 5 is divided into a radiallyouter, or primary, chamber 40 and a radially inner or secondary 38 by anannular wall 24 which, in the disclosed embodiment, is formed from anysuitable fluid-tight sheet material resistant to corrosion and to theoperating temperatures involved and having good heatconductivity, suchas suitable copper alloy or steel sheet. Wall 24 is preferably formedwith accordion pleats as shown or with other suitable convolutions, inorder to maximize the total area of heat transfer between its two sides.The axial ends of the tubular wall 24 are suitably attached to theannular inserts 27 and 28 of the rotor end walls by welding or brazing.Thus the primary and secondary chambers are completely sealed from eachother. The primary or outer chamber 40, moreover, is itself completelysealed in operation. Chamber 40, after evacuation of air therefrom, ispartly filled with a body of primary heat transfer liquid, which may bewater or other suitable liquid vaporizable and chemically stable underthe pressure and temperature conditions of operation as will be laterdescribed. The inner, secondary chamher 38 contains in operation a bodyof liquid which constitutes both a secondary heattransfer fluid and alubricant for the moving parts of the engine, specifically forlubricating the stator and rotor bearings and gearings, which liquid maybe any suitable and usual grade of oil. The oil in the secondary chamber38 inwardly of annular separating wall 24 is connected with an externalflow circuit, schematically shown, through passages as follows:

An oil inlet tube 30 is externally connected to a source of pressurelubricant such as an oil pump 86, extends through one of the statorhubs, the left one as shown, and connects internally with an annulargroove 31 formed in the inner bearing surface of the stator, thereby tolubricate said bearing. Groove 31 is connected by a short radial ductdrilled into the engine shaft 3 with a longitudinal bore 32 formed insaid shaft, which bore is shown plugged at 32A at its axially outer end,and as extending in the other or inward axial direction over a majorpart of the axial length of eccenter 4. A number of axially spacedradial holes 33 drilled through eccenter 4 discharge at their outer endsinto the secondary chamber 38. A second longitudinal bore 35 is drilledinto shaft 3 and eccenter 4 from the opposite direction and is pluggedat 35A, the bores 33 and 35 being transversely spaced from each other ina plane coinciding with the common diametric plane of shaft 3 andeccenter 4 as is visible in FIG. 1. Bore 35 is connected with thesecondary chamber 38 at the side of eccenter 4 diametrically oppositethe side containing the radial drill holes 33, by way of a similarseries of axially spaced radial drill holes 34. Bore 35 is connected bya short radial passage with an annular groove 36 formed in thecorresponding stator hub similar to inlet groove 31, and connecting withgroove 36 is an oil outlet tube 37 connected to the external lubricantsystem, specifically an air radiator or other oil cooler 88 as hereshown, from which the cool oil is returned to pump 86. The intake anddischarge grooves or chambers 31 and 36 are connected by way of narrowchannels or passages, as schematically indicated, to deliver oil tosuitable points of the bearing surfaces of the assembly.

It will be observed that in view of the relative arrangement of thelongitudinal oil conduits 32 and 35 in the eccentric 4, the radial ducts33 through which oil is discharged from conduit 32 into the secondarychamber 38 are longer than the radial ducts 34 through which oil isreturned from said secondary chamber into the conduit 35. There is thuscreated in operation a net differential centrifugal oil pressure whichaids the circulation of the oil around the secondary cooling circuit.

The primary chamber 40 defined between the outer surface of theaccordion-pleated wall 24 and the inner surface of rotor shell 14 ispartly filled with a vaporizable liquid which may be permanently sealedwithin the cavity of chamber 40. The primary liquid may be water with asuitable proportion of antifreeze agent mixed with it, or it may be anyother liquid having a suitable boiling temperature, preferably withinthe range from to C. at ordinary pressure. Thus, ethyl alcohol,trichloroethylene and tetrachloroethylene are examples of liquids usualsas the primary cooling fluid of the invention instead of or in additionto water. The primary chamber 40 is preferably filled with liquid tomore than half its total capacity, so that most of the inner surface ofseparator wall 24 is at all times contacted by said liquid even at restand at idling rates of the motor. Suitable deflecting baflle means, notshown, may be provided in the chamber 40 in a generally conventionalmanner to ensure that a substantial contact area of the liquid with theheated surfaces of the rotor shell is present even at low engine rates.

In the illustrated embodiment, the four-stroke operating cycle of theengine shown is such that in each of the five stator chambers 10, theintake and exhaust strokes occur when the rotor lobes 7 and 9,respectively, are seated with the stator lobe under consideration,whereas the compression and expansion strokes take place as the rotorlobes 6 and 8, respectively, are thus seated. Hence, as earlier noted,the lobes 6 and 8 are exposed to maximum heating, and the inner surfacesof these rotor lobes are, accordingly, shown as being provided withVapotron-type heat dissipating projections 39 in the form of parallelribs and grooves extending in a circumferential direction over the innersurface of each lobe. The ribs and grooves are shown with a triangularcross sectional contour similar to that disclosed in the applicantsco-pending US. patent application Ser. No. 512,090, filed Dec. 7, 1965.Alternatively they may be constructed as disclosed in applicants US.Patent 3,235,004. As disclosed in said patent and said patentapplication, the heatdissipating projections 39 have as their generalfunction to establish and maintain stable temperature gradients over thesurfaces thereof in contact with the vapour in the primary chamber 40,whereby higher surface temperatures can be safely reached without anydanger of local destruction or burnout of the metal.

Before considering the detailed mechanism of the evaporation heattransfer process occurring in the primary chamber 40 from the ridgedinner surface of rotor 5 to the boiling fluid in the chamber, it isimportant to consider the forces that act on the body of fluid duringthe operation of the engine, since those forces govern the distributionof the liquid and vapour phases in relation to the heat transfersurfaces at any time. There are different types of force in action, thefirst being gravity which is quite negligible in regard to the othersand will be disregarded, and the others being inertial forces created bythe movement of the rotor. As earlier indicated, the rotor describes acomplex trochoidal type of motion with respect to the stator, and itsmovement can be broken down into a component of rotation due to therotation of shaft 3 about its center axis C, and a component oftranslation due to the circular displacement of the eccentric 4 aboutthe shaft axis C. The first, rotational, component is considerably thelower of the two, and is comparatively unimportant; its effects will bebriefly referred to later. By far the principal forces acting on theprimary liquid are the inertial forces created by the second ortranslatory component of rotor motion which produces in effect, a radialreciprocation of the rotor relative to the axis of the stator. Due tothis radial reciprocation, the primary liquid sealed in chamber 40 issubjected to high radial accelerations towards and away from the statoraxis, as a result of which said liquid is alternately forced outwardsagainst the ridged rotor surface 17, and inwards against theaccordion-pleated surface of the annular separator wall 24. Betweenthese two conditions, there occurs an intermediate phase during whichthere is a mixture of liquid and vapour substantially completely fillingthe annular primary chamber and contacting both its outer (rotor)surface and its inner (separator 24) surface. While it is readilyunderstood that the three main phases or conditions just indicated mergecontinuously with one another during operation of the engine, it isconvenient to consider the three typical phases separately and they willthus be described with reference to FIGS. 35.

FIG. 3 illustrates the condition that obtains as the rotor lobe 6 istravelling radially inwards towards the stator axis as indicated by thearrow, so that the liquid phase 40A is forced into contact with theridged inner surface of said lobe, while vapour is contacting the outersurface of the pleated separator 24. At this time there is intense localboiling and vaporization of the liquid along the sides of the ridges 39and particularly in the bottoms of the grooves 42 between them. Thevapour that is thus formed is discharged radially inwards by the highcentripetal pressure generated by the radial inward rotor movement atthis time, and is forced through the annular body of liquid 40A tocollect in the inner vapour space indicated at 41, where the vapourpartially condenses in contact with the relatively cool surface ofseparator 24. During this phase, therefore, intense dissipation of heatfrom the hot rotor lobe 6 to the primary liquid 40A takes place.

As the rotor lobe 6 reverses its movement (see FIG. 5) and starts movingradially outward, the liquid 40A is forced through the vapour 41 thathas collected in the radially inward region during the preceding phaseby the strong inertial forces and an intensely active mixing actionoccurs, producing what is effectively a liquid-vapour emulsion 45filling the chamber, as illustrated in FIG. 5. The vapour condenses inthe midst of the liquid considerably cooler than it, and the averagetemperature of the body of liquid 40A rises.

The third phase, shown in FIG. 4, sets in as the rotor lobe 6 movesradially outward, so that the body of liquid 40A is now pressed againstthe pleated separating wall 24. It is chiefly during this phase that theheat accumulated by the primary liquid during the first two phases istransferred through separator 24 to the secondary liquid, the oil, inthe secondary chamber 38 forming part of the lubricating circuit earlierdescribed.

In this third phase of the heat-transfer cycle, it will be noted thatthe heat-dissipating formations 39 of the rotor lobe 6 are exposed tovapour, so that their ten pcrature rises due to the heat flux appliedthereto from the outer surface of the rotor lobe. This temperature riseoccurs mainly at the grooves 39 or roots of the protuberances, whereasthe apices 44 are heated at a substantially slower rate, a stabletemperature gradient being maintained continuously along the sides ofthe protuberances 39. Hence, at the termination of this phase when therotor will start moving inward again to recommence the first phase shownin FIG. 3, the liquid forced against the ridged wall is able to resumeimmediate contact with the relatively cool apices of the protuberancesand the normal evaporation heat transfer operation earlier described canat once be resumed, with the continuous stable temperature gradientsbeing present from the cool tips to the hot roots of the protuberances,which may be carried to temperatures well above the so-called criticalboiling temperature (about C. for water at ordinary pressure), which itwould not be permissible to exceed in the absence of the protuberanceowing to the danger of burn-out, as explained in the applicants earlierpatents and publications referred to.

In order to ensure that the protuberances 39, while exposed to thevapour and out of contact with liquid during the third phase of thecycle, will not rise to an excessively high temperature before theyrecontact the liquid as just described, it is preferred according to theinvention that said protuberances are made relatively massive so as tohave substantial thermal inertia.

It will be seen from the above detailed description of the cyclicevaporation heat transfer process occurring in the rotary-piston enginedescribed, that the over-all heat transfer from the outer surface of thehot rotor lobe 6 by way of the primary boiling fluid 40A to thesecondary lubricating liquid, is extremely efficient. The efficiency isdue in part to the relatively high temperatures attainable at thehottest points, i.e. in the grooves, of the ridged rotor surface withoutdanger of burnout, and is also due largely to the intensely activemixing and turbulence of the vapour and liquid during the intermediate(emulsion) phase described with reference to FIG. 5 of the describedheat transfer mechanism. As a result there is only a comparatively smalltemperature drop between the primary heat transfer wall (the ridged Wallof rotor lobe 6) and the secondary heat transfer wall (separator 24) andhence the secondary coolant-lubricant liquid, regardless of the distancebetween the two walls and hence the size of the engine, the heat beingso to speak transported by the body of primary liquid from the primarywall to the secondary Wall by the inertial forces inherent to theoperation of the engine. The over-all heat resistance of the twocascaded heat exchange means remains very low even when the flux densityof the heat to be dissipated is high, so that it becomes possible to useheat flux densities several times higher than the highest valuesheretofore usable in engines of the type described.

As earlier mentioned, the circular component of rotor movement also hasan action on the heat transfer cycle. This component acts to impart tothe annular body of liquid in the primary chamber a turbulent bodilyrotation with respect to the rotor. The resulting centrifugal forcesonly have a minor modifying effect on the cyclic process describedabove. However, the bodily rotation of the liquid annulus is beneficialin that it serves to equalize the temperatures between the four lobes ofthe rotor which, as earlier indicated, are unequally loaded thermally inthe 4- stroke embodiment described. The effect is especially markedduring periods of acceleration and deceleration due to the increasedturbulence at such times. This improves the uniformity of the heattransfer and, additionally, tends to equalize the thermal expansioneffects to which the rotor is exposed.

In the embodiment of the invention just described, additional cooling isused for the stator, and for this purpose the peripheral chamber 12earlier referred to, is filled with a vaporizable liquid for instancewater. The outer surfaces of each of the five combustion chambers 16 ispartly formed with heat dissipating protuberances 13, such as ridges andgrooves, which may be formed similar to the protuberances 39. Thechamber 12 is prefeiably connected with an external circulatory systemby way of an inlet and an outlet not shown. Alternatively it may besealed and cooled by the surrounding air, or a secondary cooling liquid.

Returning to the rotor cooling system and the cyclic heat transferprocess earlier described with reference to FIGS. 3-5, it will beobserved that the pressure within the sealed primary chamber 40 issubjected to large and rapid variations during each cycle, the pressurerising during the phase shown in FIG. 3 and dropping in the phase ofFIG. 4, and the pressure being especially low in the emulsion phase ofFIG. 5. The average pressure, however, will clearly increase withincreasing heat dissipation in the engine, due to the increase inpressure of the vapour produced. Now, any increase in operating pressurewill improve the efficiency of each of the heat transfer processesinvolved, including the efiiciency of the vapori zation heat transfer,that of the condensation heat transfer and that of the conduction heattransfer through separator 24. In other words, the efficiency of theheat transfer increases as the operating rate of the engine isincreased. This constitutes in effect a natural feedback action which isvery desirable.

Some numerical values for the temperatures and pressures present invarious parts of the rotor cooling system of a rotary-piston engineconstructed in accordance with the invention, will now be given by wayof example.

The oil in the secondary cooling system is pumped with a flow rate suchthat its temperature rise on contacting the separator 24 does not exceed20 C. at the maximum engine operating rates, and is passed through anexternal cooling system such as a radiator 88 whereby the averagetemperature of the oil is held at about 70 C. The separating orsecondary heat transfer wall 24 is arranged to transfer a heat flux ofabout watts per sq. centimeter with a temperature drop across said wallnot exceeding 30 C. It is noted that this conduction rate is readilyachieved in view of the great tubulence of the primary liquid contactingthe wall. The primary liquid, therefore, has an average temperature ofabout 70+30=l00 C. The highest pressure reached by the vapour in thelimited vapour space provided in the primary chamber 40, which vapour isgenerated in the grooves of the protuberances 39 as earlier described,is found to be in the range of from 2 to 4 atmospheres, whichcorresponds to a boiling temperature of about 120 C. It is here notedthat said maximum vapour pressure is limited only by the condensationsimultaneously occurring in the liquid. The condensation produced bymixtures of the vapour with liquid in a turbulent emulsion of the kindpresent during the transitional phase of the heat transfer cyclepreviously described herein (FIG. 5 is very efficient and its efficiencyrises rapidly as the temperature difference between the vapour andliquid phases increases. Experience has shown that a temperaturedifference of only C., as present between the 100 C. primary liquid andits vapour at 120 C., is amply sufiicient to ensure more than therequired condensation capacity.

It should be observed that the 2 to 4 atmospheres pressure range in theprimary chamber 40 is substantially of the same order of magnitude asthe oil pressures normally used in engine lubricating systems. Thepressures on both sides of the separating wall 24 are therefore aboutthe same, and this simplifies the construction of this wall e.g. fromcorrugated metal sheet having good heat transfer characteristics, sincesaid wall will not be subjected in operation to high mechanicalstresses.

Turning now to the primary heat transfer wall constituted by the rotorshell 51, it is important that the heat dissipating extensions 39 beproportioned in accordance, with the teachings of the prior patent orapplication identified above, in order to provide full assurance againstthe danger of burnout of the rotor. As described in application Ser. No.512,090, filed Dec. 7, 1965, the dimensions of the extensions 39 can bedetermined from the heat conductivity of the metal constituting saidwall, the nominal value of the heat flux density present through thewall at the maximum rate of engine operation, and certain other physicalproperties of the primary liquid used. The parameters available make itpossible to predetermine said dimensions for a good economy and maximumperformance in each particular instance of use. The rotor wall 5 whenthus constructed will be capable of transferring heat flux densities ofseveral hundred watts per sq. em while having an average temperature inthe groove 43'between the extensions 39 only 20 or 30 C. higher than thetemperature of the vapour generated in said grooves.

From the temperature values given above for both the secondary and theprimary heat transfer assemblies described, it can be seen that themaximum temperature at the primary heat transfer surface 6 issubstantially less than 200 C. Thus the outer surface of said rotor wallwhich is exposed to the heated gases does not greatly exceed about 200C. even in the presence of an intense heat flux. Tests have confirmedthese results, and it is evident that greatly reduced operatingtemperatures obtainable with the cooling system of the invention willcorrespondingly improve the operating conditions of the engine,including the efficiency ratio, the lubrication of the relatively movingsurfaces and generally the performance and service life of the engine.

An additional advantageous feature of the construction described is thatthe piston rings or seals such as 17 are exposed to very active coolingowing to their proximity to the heat-dissipating formations 39. Athigher operating loads; the evaporation of the primary coolant liquid isnot restricted to said heat dissipating formations but extends also tothe adjacent surface areas 46 of the rotor and flanges, thereby assuringenergetic cooling of said piston rings.

In FIG. 6 the invention is shown as embodied in a reciprocatory piston,e.g. of a marine engine or the like. The piston comprises a skirt 51 andan endwall 47, shown outwardly concave, formed as an integral casting,and the inner surface of the endwall 47 being formed with heatdissipating extensions of the type disclosed e.g. in the applicantslast-identified patent application. Conventional piston rings 48 areseated in grooves 49 formed in the upper peripheral surface region ofthe skirt. A plugging disk 50 is sealingly secured, as by welding,across the inner bore of the skirt 51 not far from its open and, so asto define between the disk 50 and the ridged surface of the endwall 47 asealed space or chamber. This constitutes the primary chamber of thetwo-stage cooling system of the invention and is partly filled with asuitable vaporizable liquid, which may be water or any of the otherliquids earlier mentioned as suitable.

The sealing disk 50 is formed with aligned bearings 52 and 53 in whichthe cross'pin 54 of the connecting rod 55 is journalled. The secondarycooling system in this embodiment comprises a coil pipe 57 mounted inthe primary chamber on the inner surface of disk 50 and having an inletend connected to a passage 58 in the crosspin 54 and thence a passage 59in the connecting rod 55. The free end of the coil 57 is connected byway of a calibrated orifice 60 with the interior of a sealed chamberdefined by a dome 61 upstanding from the disk 50. The coil 57 is mountedcoaxially around this dome and is supported in place by spacers such as62. The secondary circuit may comprise the usual accessories not shownincluding an oil pump for circulating the oil through passage means inthe crankshaft (not shown) and by way of the crank bearing, which it mayserve to lubricate in passing, up the passages 59 and 5-8 through thecoil 57 and calibrated orifice 60, which serves to determine the flowrate of the oil, into outlet chamber defined in dome 61. From thischamber the oil may flow out by suitable outlet means not shown,directly into the crankcase, preferably lubricating the crosspin hearing54 on its way out.

The plugging disk 50 may have a scalable filling plug,

1 1 not shown, extending through it for introducing the primary liquid(e.g. water) into the primary chamber within the piston.

The operation of this embodiment will be easily understood from theexplanations given in connection with the first embodiment of theinvention, and will not be described in detail. As in the case of therotary-piston engine, the heat transfer process proceeds in a cycledetermined largely by the movements of the body of primary liquid withinthe piston chamber, under the inertial forces created by thereciprocation of the piston. The heat generated in the compressionchamber of the engine (not shown) and applied to endwall 47 of thepiston, is thus carried away from the ridged inner wall 56 provided withthe Vapotron heat dissipating extensions 39, and is transferred throughthe wall of the coil 57 to the lubricant, to be dissipated thereby inthe external lubricant system.

It will be understood that the cylinder heads, not shown, of thereciprocatory engine including the piston just described, may if desiredbe provided with evaporation cooling means generally similar to themeans described and shown for the corresponding parts of therotary-piston engine of FIGS. 1-2, including if desire the Vapotron likeheat dissipating extensions shown at 13 in FIG. 2.

The heat dissipating formations referred to in the last paragraph and atother points of the disclosure, in particular the formations 39 providedon the inner surfaces of the rotor lobes in FIGS. 1 and 2 and the innersurface of the piston and wall in FIG. 6, may assume a wide variety offorms. Preferably they are of either of the types disclosed in theapplicants prior US. Patent 3,235,- 004 or in the applicants copendingpatent application Ser. No. 512,090, filed Dec. 7, 1965.

According to the said patent, the extensions or protuberances andintervening channels are so dimensioned as to satisfy substantially therelations wherein d represents the average transverse Width of aninterprotuberance channel, a the average transverse width of aprotuberance between adjacent channels, the heat conductivity factor ofthe wall material, and m a numerical coefficient within the range fromabout 0.7 to about 1.8, when a and b are expressed in centimeters and cin watts transmitted heat per centimeter and per degree centigrade.

According to the said copending application, the extensions orprotuberances and intervening channels are so dimensioned as to satisfysubstantially the relations wherein 17 represents the height of aprotuberance, .r and s the base area and total side surface area ofprotuberance respectively, 0 the heat conductivity coefficient of thewall material, q the critical value of heat flux density of said liquidat the operating pressure, 0 a specified temperature drop from the baseto apex of a proturberance, 1 the maximum specified value of heat fluxdensity per unit area of the heat input surface, k a numerical safetyfactor selectable over the range from 1 to 2, and p a numericalefiiciency factor selectable over the range from 0.8 to 1.6.

What I claim is:

1. In a machine including a moving member and wherein heat is generatedadjacent a wall surface of said member, a cooling system for said membercomprising:

means defining a sealed primary enclosure in said member adjacent tosaid wall surface thereof;

a body of a primary coolant liquid that is vaporizable in the range oftemperature permissible for said member, sealed in the primaryenclosure;

means defining a secondary enclosure in said member in a region spacedfrom said wall surface;

means defining a flow system for a secondary coolant fluid connectedwith the secondary enclosure to circulate said fluid therethrough; and aheat exchanging and separating wall extending between said primary andsecondary enclosures, said heat exchanging and separating wall havinggood heat conducting characteristics and being shaped as a heatexchanger between two moving liquids;

whereby said primary liquid will abstract heat from said wall surface,and transfer the abstracted heat through the separating wall to thesecondary coolant fluid.

2. The combination defined in claim 1, wherein the surface of said heatexchanging wall directed into the enclosure is provided withheat-dissipating formations, said heat-dissipating formations comprisingprotuberances separated by channels and so dimensioned as to establishstable temperature gradients over the width of the wall directed towardsthe primary enclosure.

3. The combination defined in claim 2, wherein said extensions aredimensioned in substantial accordance with the relations wherein drepresents the average transverse width of an interprotuberance channel,a the average transverse width of a protuberance between adjacentchannels, 0 the heat conductivity factor of the wall material, and m anumerical coefiicient within the range from about 0.7 to about 1.8, whena and b are expressed in centimeters and c in watts transmitted heat percentimeter and per degree centigrade.

4. The combination defined in claim 2, wherein said etxensions aredimensioned in accordance with the relations wherein b represents theheight of a protuberance, s and s the base area and total side surfacearea of a protuberance respectively, 0 the heat conductivity coefficientof the wall material, q the critical value of heat flux density of saidliquid at the operating pressure, 0 a specified temperature drop fromthe base to the apex of a protuberance, b the maximum specified value ofheat flux density per unit area of the heat input surface, k a numericalsafety factor selectable over the range from 1 to 2, and p a numericalefficiency factor selectable over the range from 0.8 to 1.6.

5. The combination defined in claim 1, wherein said heat exchanging andseparating wall is formed with convolutions for increasing the surfacearea of its contact with said primary liquid and secondary fluid.

6. The combination defined in claim 1, wherein the primary liquid has aboiling temperature in the range of from to C. at ordinary pressure andis chemically stable in said range.

7. In a machine including a moving member whose motion has a componentsubject to a periodically varying acceleration and wherein heat isgenerated adjacent a wall surface of said member, a cooling system forsaid member comprising:

means defining a sealed primary enclosure in said member adjacent tosaid wall surface thereof;

a baby of a primary coolant liquid that is vaporizable in the range oftemperatures permissible for said member, sealed in the primaryenclosure;

means defining a secondary enclosure in said member in a region spacedfrom said wall surface;

13 means defining a flow system for a secondary coolant fluid connectedwith the secondary enclosure to circulate said fluid therethrough; and

a heat exchanging and separating wall extending between said primary andsecondary enclosures in a general direction transverse to said componentof motion of said member, said heat exchanging and separating wallhaving good heat conducting characteristics and being shaped as a heatexchanger between two moving liquids whereby said sealed body of primaryliquid will be intermittently thrown against said separating wall by theinertial forces created by said periodically varying acceleration topromote heat transfer from said primary cooling liquid to said secondarycoolant fluid.

8. The combination defined in claim 7, wherein the surface of said heatexchanging and separating wall directed into the primary enclosure isprovided with heatdissipating formations.

9. The combination defined in claim 7, wherein said secondary coolantfluid comprises a lubricant liquid for said machine.

10. The combination defined in claim 7, wherein said machine is arotary-piston machine of the type including a multi-lobed rotordescribed a trochoidal motion relative to a stator so as to definevariable-volume working chambers therein, and wherein said rotorcomprises said moving member.

11. The combination defined in claim 10, wherein said machine includes ashaft mounted for rotation in the stator, an eccenter secured on theshaft, and means mounting said rotor for slidable motion on theeccenter, and wherein said separating wall comprises an annular wallsurrounding the eccenter and defining an annular outer primary enclosureoutwardly limited by the inner wall surfaces of the rotor lobes, and anannular inner 14 secondary enclosure inwardly limited by the eccenter.12. The combination defined in claim 11, wherein said separating wall isin the form of an accordion pleated sleeve.

13. The combination defined in claim 11, including passage means formedin said shaft and said eccenter for connecting the secondary enclosurewith said flow systern.

14. The combination defined in claim 11, wherein an inner surface of atleast one of said rotor lobes is formed with heat dissipating formationsin the form of protuberances projecting into said primary enclosure andchannels separating said protuberances, and so dimensioned as toestablish stable temperature gradients over the side of said walldirected towards the primary enclosure.

15. The combination defined in claim 7, wherein said machine is areciprocatory-piston machine, and the piston constitutes said movingmember.

16. The combination defined in claim 15, including means defining a wallsealingly extending across said piston and spaced from the piston endwall to define said primary enclosure having said body of primarycoolant liquid sealed therein.

17. The combination defined in claim 16, including a tubular coilpositioned within said primary enclosure and defining said secondaryenclosure connected with the secondary fluid flow system.

References Cited UNITED STATES PATENTS 2,175,265 10/1936 Johnson.3,176,915 4/1965 Bentele 123-8 X 3,303,829 2/1967 Castelet 123-8 RALPHD. BLAKESLEE, Primary Examiner.

